Technical Papers. 31st Annual Meeting. International Institute of Ammonia Refrigeration. March 22 25, 2009

Transcription

1 Technical Papers 31st Annual Meeting International Institute of Ammonia Refrigeration March 22 25, Industrial Refrigeration Conference & Exhibition The Hyatt Regency Dallas, Texas

2 ACKNOWLEDGEMENT The success of the 31st Annual Meeting of the International Institute of Ammonia Refrigeration is due to the quality of the technical papers in this volume and the labor of its authors. IIAR expresses its deep appreciation to the authors, reviewers and editors for their contributions to the ammonia refrigeration industry. Board of Directors, International Institute of Ammonia Refrigeration ABOUT THIS VOLUME IIAR Technical Papers are subjected to rigorous technical peer review. The views expressed in the papers in this volume are those of the authors, not the International Institute of Ammonia Refrigeration. They are not official positions of the Institute and are not officially endorsed International Institute of Ammonia Refrigeration 1110 North Glebe Road Suite 250 Arlington, VA (voice) (fax) Industrial Refrigeration Conference & Exhibition The Hyatt Regency Dallas, Texas

3 Technical Paper #7 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Stefan S. Jensen, B.Sc.Eng. MIEAust Scantec Refrigeration Technologies Pty. Ltd. Brisbane, Queensland, Australia Abstract In view of the very large range of evaporator coil geometries, coil material combinations, coil defrost methods and circuiting options available on the market, the industrial refrigeration contractor often faces difficulties deciding which evaporator design to use for a certain application. Often selections are based on rules of thumb, e.g., an allowance of a certain number of square feet of coil surface area per pound of product. Considering the fact that coil material selections, choice of refrigerant, coil circuiting and coil geometry can influence heat transmission coefficients (u-values) by up to a factor of three or more, such an approach can lead to very poor results in practice. Other issues are relative humidity and dehydration of the refrigerated space. There are thousands of practical applications where the quality and shelf life of the products stored or chilled are directly influenced by the equilibrium relative humidity inside the room in question. Yet often extended surface air coolers are selected with geometries to suit the manufacturer and not the application. Frequently, the result is a squat coil when the coil should have been shallow and with large face area. The performance impact of fouling both on the inside, but also on the outside surfaces of cooling coils is often not highlighted sufficiently by manufacturers. Extended surface air cooling coils featuring comparatively high-heat transmission coefficients (u-values) in clean condition generally display a more rapid performance deterioration as a function of increasing fouling than an equivalent coil with lower heat transmission coefficient in clean condition. It is often not in the commercial interest of coil manufacturers to disclose this information, yet it can be of crucial importance to the refrigeration plant designer. It is commonly known that cooling coil performance increases as a function of increasing face velocity. What is not normally available from manufacturers is optimization of face velocities for maximum plant energy efficiency. This is the contractor s problem and the tools that the contractor needs to perform this task are often not readily available. This paper will show a range of practical performance comparisons between various coil geometries, coil materials, circuiting options, refrigerant choices, flow patterns and refrigerant feed options. These performance comparisons will have one thing in common namely the software used to perform them. The comparisons therefore represent relative information valuable to the contractor because it is information not readily available from extended surface cooling coil manufacturers under normal circumstances IIAR Ammonia Refrigeration Conference & Exhibition, Dallas Texas IIAR

4

5 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Introduction In modern industrial refrigeration applications, the extended surface air cooler is one of the most common components. In terms of importance for the success of an installation, it ranks at the level of compressors, condensers and control systems. Although many software packages exist for the selection and rating of air coolers, these are all designed for certain proprietary geometries and manufacturing ranges. They are therefore not readily suitable for relative performance comparisons between one manufacturer and the other. The expertise of industrial refrigeration contractors relates to the ability to successfully combine key plant components to form a functional refrigeration system to suit a certain technical application. The selection of extended surface air coolers can be a complex matter particularly if the outcome specified goes beyond simple coil performance issues and extends into factors such as product quality, weight loss, defrost frequency, overall energy efficiency, noise pollution, hygiene, temperature and air flow uniformity within a facility, freezing times, part load behavior and other similar performance outcomes for the system as a whole. Not all extended surface air cooler geometries and layouts are suited optimally to all applications. To some this may be a statement of the obvious and to some extent it is. The statement does not refer to basics like selecting the correct fin spacing for a freezing application. The statement makes the point that within one particular plant it may be necessary to employ a wide range of cooler geometries and layouts to achieve the optimal combination of components suitable for the application. This may mean mixing coil manufacturers within the same system. The marketing emphasis of cooler manufacturers is generally that their particular product offered covers the widest range of applications. This is true for those manufacturers who offer a wide range of geometries and construction materials. Those manufacturers who for one reason or another limit the offering to one or Technical Paper #7 IIAR

6 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas very few coil geometries cannot cover as wide a set of applications. This paper will examine the contractor s perspective when it comes to selecting optimal extended surface air coolers for a range of applications. The Extended Surface Air Cooler The typical air cooler type which is the topic of this paper is shown in principle (Figure 1). It consists of a tube bundle fitted with plate fins; air is forced through the fin side, refrigerant flows through the tubes. The tube pattern can be either square or triangular; the tube diameters, fin thickness, fin design, tube centre distances, circuiting, face velocity, refrigerant, refrigerant feed and material selections are variables. This paper will not cover the issues of turbulators in tubes or wavy fins nor will it describe any other fluid on the fin side than humid air at atmospheric pressure. Finally, the paper will generally be limited to the presentation of results to maximize clarity. The air cooler performance modeling methodology used for the various comparisons in this paper is described in earlier papers by the author [Reference 1], [Reference 2]. Impact of Coil Geometry on Overall Heat Transmission (U-Value) The term, coil geometry, refers in this context to the tube pattern and the centre distances between the tubes. Cooling coil layout refers to the way a particular cooler of a particular geometry is laid out, i.e., rows high, rows deep, circuiting, finned length, fin spacing, etc. In Figure 2 is shown a range of hot dipped galvanized steel coils of different geometries. By using the same material throughout the range, the variation in heat transmission coefficients from 34.6 to 18.9 W/m 2 K (6.09 to 3.33 Btu/h ft 2 F) is the 4 IIAR 2009 Technical Paper #7

7 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective result of variations in coil geometry only. All coils shown use 0.3 mm (0.0118") thick fin material with a mm ( ") thick layer of zinc. Impact of Materials of Construction and Fin Thickness Cooling coil performances are influenced by materials of construction and fin thickness. This is demonstrated in Figure 3 where the coil geometry has been kept constant, but materials and fin thicknesses have been varied to isolate the impact of these variables. The variation in heat transmission coefficients from 13.2 to 37.5 W/ m 2 K (2.32 to 6.60 Btu/h ft 2 F) is a direct result of the changes to thermal conductivity of materials and fin efficiencies. Circuiting Circuiting is how the refrigerant is piped through the cooling coil. A coil can be piped for physical parallel flow, physical counter flow, thermodynamic counter flow, a mixture of physical parallel flow and counter flow and cross flow between air and refrigerant. In addition, the circuits can be made long or short as dictated by the tube diameter, refrigerant temperature, heat flux and refrigerant properties. If the application requires relatively short circuits and the coil in question is large, then it is often necessary to provide the coil with multiple headers or manifolds because there is a manufacturing limitation with respect to the number of circuits which can enter/ leave a header. This makes the valve station more complicated and costly and is generally not in the interest of the contractor. The way the circuits run in relation to the force of gravity is dictated by a number of factors. These include oil drainage, the type of refrigerant feed, orientation of the coil, heat flux, type of defrost, thermostatic expansion valve sensor location and general refrigerant side thermodynamics. Circuiting errors are most forgiving in liquid Technical Paper #7 IIAR

8 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas overfeed applications and potentially most disastrous in dry expansion feed cooling coils. The greater the turbulence on the refrigerant side, the greater the surface heat transfer and the greater the pressure drop. The greater the refrigerant side pressure drop for a volatile refrigerant, the greater the refrigerant temperature drop as the fluid passes through the coil. The greater the refrigerant temperature drop, the lower the logarithmic mean temperature difference (LMTD) all other things being equal. Lower LMTD translates into reduced coil capacity. This scenario presents an optimization problem between achieving good heat transfer on the tube side of the heat exchanger without jeopardizing LMTD. A relatively high refrigerant side temperature drop has the capacity to reduce LMTD to such an extent that any gain achieved by improved heat transfer is more than eroded by a reduction in temperature difference hence reducing overall coil capacity. It is generally not possible to circuit any cooling coil in an optimal manner for the entire operating range that the coil will encounter during its operating life. All choices of circuiting therefore represent a compromise. If the manufacturer is provided with the range of design operating conditions applicable to the coil in question then it is possible to determine the most reasonable circuiting for the operating envelope nominated. In many cases, however, the manufacturer is only provided with one operating point namely the maximum coil performance required. Particularly in dry expansion applications this can in practice lead to poor results at part load (low temperature difference) and/or at reduced face velocity [Reference 2]. By far the most common circuiting method for industrial applications is shown in Figure 4. This shows a liquid overfeed coil with horizontal headers; liquid inlet at the bottom, wet return at the top; horizontal air flow and four tube passes. In effect, this represents a cross flow arrangement between air and refrigerant. 6 IIAR 2009 Technical Paper #7

9 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective If a coil such as the one shown in Figure 4 is for a low temperature application and is both long between the end plates and comparatively tall to meet the capacity requirement, then the number of tube passes needs to reduce. This is to ensure that the refrigerant side temperature (pressure) drop does not jeopardize LMTD to such an extent that overall capacity reduces. It is generally less expensive for the coil manufacturer to then supply a coil with multiple horizontal liquid and wet return headers as opposed to having one inlet and one outlet header that accommodate all the circuits. A typical multiple header coil is shown in Figure 5. The consequence to the contractor of this header arrangement is that each evaporator segment needs to be treated like an individual evaporator in terms of valve station design. Essentially, the capital cost of the air cooler has therefore been minimized at the expense of increased valve station costs. Often the result is an overall project cost increase because the additional cost of the valve station more exceeds the saving in coil costs. An alternative to multiple horizontal headers to achieve the correct circuit length is vertical headers (Figure 6). In the great majority of cases, coil depth is less than the coil height so the need for multiple headers can be eliminated by using vertical headers. However, to ensure uniform distribution of refrigerant in a liquid overfeed situation, each circuit inlet must have an orifice. If there are no orifices, the refrigerant will take the path of least resistance and the upper circuits of the coil will not perform. This may be illustrated with a simple example. If the refrigerant side pressure drop for the coil illustrated in Figure 6 is 1K (1.8F) then this translates into a pressure drop of 3839 Pa (0.56 psi) or 0.57 m (1.86 feet) of ammonia liquid column. If the vertical liquid inlet header is 1.2 m (3.94 feet) tall, then gravity will force the liquid refrigerant through the bottom half of the evaporator circuits leaving the top circuits dry. The result will be more or less a halving of the anticipated cooling capacity. The design driving force across the circuit orifices should be greater than the height of the vertical inlet header. It follows therefore that the top orifices in the header need to be larger than the bottom orifices to account for gravity. The mass flows through circular orifices may be calculated as shown: Technical Paper #7 IIAR

10 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas m = α * A * SQR(2 * p * ) (1) For the orifice contraction coefficient, a value of 0.7 may be used. For the top orifices in the header a design pressure drop ( p) of around 0.45 to 0.6 m (1.5 to 2 feet) of liquid column is practical and will generally not yield orifices which are too small and hence susceptible to blockage. The design pressure drops measured in liquid column for the bottom orifices are the same as for the top plus the height of the header. Using larger design pressure drops will reduce the size of the orifices to such an extent that they become susceptible to impurities. For the purposes of streamlining manufacturing operations, some coil manufacturers only provide a choice between either ø3 or ø4 mm (0.12" or 0.16") orifices. In most cases these are far too large (unless the liquid overfeed ratio is very high) and do not provide the appropriate graduation in size between top and bottom of the inlet header. More appropriate orifice sizes are around ø1.7 to ø2.8 mm (0.067" to 0.11"). Even with these orifice sizes, it is often found in practice that the liquid overfeed ratio needs to be elevated to around 6 to 1 to ensure adequate refrigerant distribution to all circuits during all operating conditions. The orifices in a vertical inlet header not only restrict the flow of liquid they also restrict the flow of hot gas. This needs to be carefully considered when an evaporator is to be defrosted with hot gas. It is often found in practice that coils with vertical headers require longer time to defrost with hot gas than coils with horizontal headers where the hot gas is entered in the top wet return header hence pushing the cold liquid out of the bottom header relatively rapidly at commencement of defrost. The circuit orifices in the coil shown in Figure 6 may be eliminated by turning the coil 90 such that the air flows vertically up through the coil. In liquid overfeed applications this then enables liquid entry through the bottom header and wet return from the top thus creating physical parallel flow or thermodynamic counter flow. Simply turning the coil 90 is naturally not always physically possible, but if it is, then a circuiting problem can be converted to a circuiting advantage. This circuiting 8 IIAR 2009 Technical Paper #7

11 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective method has been used successfully in a large number of automatic tunnel freezers in Australia and New Zealand for about 50 years. There are cases where circuit orifices may be required also in coils with horizontal headers and horizontal air flow. In a situation where the coil is exposed to high ETD for example a fresh air inlet cooling coil in a food processing plant the circuits at the air inlet to the coil may boil dry. This is due to the fact that high heat flux also creates high refrigerant side pressure drop. At the air leaving side of the coil where the heat flux is less, the refrigerant side pressure drop is relatively low. Without circuit orifices the circuits at the air leaving side are then in equilibrium supplied with excess refrigerant at the expense of the circuits at air entry, which are left starved of refrigerant. The result is a significant reduction in overall coil performance (Figure 7). The impact of circuiting on coil performance is visualized in Figure 8. The number of rows high, deep, fin spacing and finned length have all been kept constant. The number of times that the refrigerant passes through the tube bundle has been varied this is referred to as the number of tube passes. The vertical axis shows heat transmission coefficients and gross coil capacity; the horizontal axis shows evaporating temperature and entering temperature difference (ETD) between air and refrigerant. The ETD is calculated by subtracting the evaporating temperature from the air on temperature of 18 C ( 0.4 F) shown. The graphs show the coil performance as a function of the number of tube passes. It is clear that there is no such thing as optimal circuiting for a wide range of operating envelopes. At 25 C ( 13 F) evaporating temperature (7K or 12.6 F ETD), 28 passes deliver maximum coil performance. At 28 C ( 18.4 F) evaporating temperature, 14 passes yield maximum performance. In a dry expansion application such as the one shown, circuiting for a comparatively low refrigerant pressure drop at full load may deliver optimal coil performance at that point. However, practical experience has also shown that a coil circuited this way is likely to become difficult to control at low ETD s and Technical Paper #7 IIAR

12 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas the result will be increasing liquid entrainment in the suction vapor leaving the coil. This is also known as liquid flood-back. It is possible to circuit both dry expansion coils and liquid overfeed coils such that the reduction in LMTD by a relatively high refrigerant side temperature (pressure) drop is minimized (Figure 9). Physical parallel flow translates into thermodynamic counter flow. Using this circuiting method will, in dry expansion applications, give rise to problems obtaining a suitable superheat signal for the expansion valve. These problems may be overcome by circuiting the last tube pass through the air inlet side (Figure 10). This circuiting method is termed reversed suction return. The difference in coil performance between reversed suction return and physical counter flow for a typical dry expansion coil is illustrated in Figure 11. Fouling Fouling is a layer of material with relatively poor thermal conductivity that has settled on the heat transfer surface of a heat exchanger. This layer has the capacity to create a thermal resistance and hence inhibit heat transfer. The result is a reduction of the overall heat transmission coefficient. In industrial extended surface air coolers, fouling can occur both on the refrigerant side and on the air side. On the refrigerant side of air coolers in ammonia plants, the most common form of fouling is oil; on the air side it is frost. There are other forms of fouling, but these are outside the scope of this paper. Not all air cooler geometries are affected the same way by a given fouling resistance. The percentage coil performance reduction as a function of the internal and external fouling resistances is shown in Figure 12 for two different coil geometries. It is evident that the performance of geometry (a) deteriorates more rapidly than that of geometry (b) as fouling increases. Due to the fact that fouling is almost unavoidable in industrial ammonia refrigeration plants, this observation can be of crucial 10 IIAR 2009 Technical Paper #7

13 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective importance to the contractor. A thin layer of oil of 0.05 mm (0.002") thickness on the refrigerant side of an ammonia air cooler is not uncommon [Reference 3]. This results in a fouling resistance of m 2 K/W ( ft 2 hr F/Btu). The impact of this fouling resistance is a reduction in heat transmission coefficient of 13.5% in the case of geometry (a). A frost layer of ~1 mm (0.04") throughout the coil represents a fouling resistance of around m 2 K/W (0.046 ft 2 hr F/Btu) [Reference 4]. The impact of this is a capacity reduction of 27.8% and 21.4% for geometries (a) and (b) respectively. Refrigerant and Refrigerant Feed A very common refrigerant feed method in industrial ammonia applications is liquid overfeed. Figure 13 shows the performance of a typical ammonia liquid overfeed coil as a function of liquid overfeed rate. Although the liquid overfeed rate has little impact on the coil performance when viewed in isolation, excess liquid overfeed rates can have a significant impact on the pressure drop in the wet return line downstream of the evaporator. An increase in wet return line pressure drop increases the ammonia pressure that the coil is exposed to at the suction connection and the result is a reduction in coil capacity. Common alternatives to liquid overfeed are gravity flooded and dry expansion feeds. The principle of gravity flooded feed is shown in Figure 14. In equilibrium, the sum of pressure gains and pressure drops in the refrigerant circuit must be zero. In Table 1 are shown the key data for a flooded evaporator with NH 3 and CO 2 refrigerants. The air flow is vertically up (Figure 14). Technical Paper #7 IIAR

14 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas Table 1. Key data for NH 3 and CO 2 flooded evaporators Refrigerant NH 3 CO 2 t E, C (F) 41.5 ( 42.7) 41.5 ( 42.7) t S, C (F) 42.4 ( 44.3) 41.8 ( 44.3) t E + P R, C (F) 39.7 ( 39.5) ( 42.63) Air on/off, C (F) 35.0/ / 38.2 ( 31.0/ 36.0) ( 31.0/ 36.8) Relative humidity on/off, % 80/93 80/93 Liquid overfeed rate 5.2 to to 1 Total cooling capacity, kw (TR) (35.6) (40.8) Overall heat transmission coefficient, service, W/m 2 K (Btu/h ft 2 F) 31.6 (5.56) 32.1 (5.65) d 1, m (inches) (4.03) (3.07) d 2, m (inches) (3.07) (2.47) d 3, m (inches) (5.05) (3.07) For both refrigerants the evaporator coils and surge drum heights H L are identical. The interconnecting refrigerant pipe line diameters have been selected in accordance with good design practice. The impact of the difference between pressure/ temperature gradients for the two refrigerants is evident in Table 1. The practical consequence of the greater temperature change of NH 3 is reduced LMTD across the NH 3 evaporator and hence lower performance. Dry expansion refrigerant feed is very common in commercial and small to medium size industrial applications. Some characteristics of dry expansion feed are less favorable than those of liquid overfeed. This becomes evident at evaporator part-load i.e., at reduced face velocity, reduced ETD and when frost accumulates on the coil. A comparison between dry expansion and liquid overfeed evaporators is provided (Figure 15). The leveling or declining capacity at high ETD s for dry expansion feed coils is evidence that the refrigerant side pressure (temperature) drop becomes so 12 IIAR 2009 Technical Paper #7

15 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective high that it starts to jeopardize LMTD and hence overall capacity. The liquid overfeed coil capacity continues to climb when exposed to the same operating condition. Air Pressure Drop The air pressure drop through a finned air cooler is a function of air cooler geometry and face velocity. Figure 16 provides some typical air pressure drop comparisons for a number of geometries at identical face velocities. The geometries, coil layouts and operating conditions are identical to those shown in Figure 2. The air pressure drop increases with increasing face velocity. The rate of increase may be approximated as the square of the face velocity increase. Energy Optimization The gross air cooler capacity increases with increasing face velocity. Gross and net air cooler capacities are in this context defined as cooler capacities without and with correction for fan power respectively. To force a certain quantity of air through a cooling coil with a certain air pressure drop requires a certain amount of fan power. In the majority of refrigeration, chilling and freezing applications, the fan power is converted to heat, which enters the refrigerated space and in turn needs to be removed again by the air cooler. In most cases it is therefore only the net cooler capacity (gross capacity minus fan heat) that is available for refrigerating the space in question. Plotting net cooler capacity as a function of face velocity using the ETD as the parameter reveals that the most energy efficient face velocity is not constant, but varies depending on ETD, (Figure 17). For relatively high ETD s, it is more energy efficient to operate with relatively high face velocities and vice versa. These data sets may be generated for any combination of air cooler and fans. The data set is useful Technical Paper #7 IIAR

16 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas for producing a look-up table used in the electronic control system for an evaporator where the fans are controlled by means of a variable frequency drive (VFD). The line shown through the optimum operating points is of significance for the establishment of this look-up table. Measurement of refrigerant and air entering temperature defines ETD and this in turn enables the control system to automatically set the most energy efficient fan speed. Humidity The minimization of chilling weight loss is of vital importance during the cooling of most perishable goods where the cooling process uses air as the secondary refrigerant surrounding the goods (Figure 18). In addition, the maintenance of high relative air humidity in spaces where perishable goods such as leafy vegetables, flowers and fruits are stored is equally important to maximize shelf life. The importance of cooler geometry and cooler layout is often overlooked in this context. A typical average refrigeration load in a batch type beef carcass chiller with a holding capacity of 24,000 kg (52,911 lbs) and a design chilling cycle time of 18 hours may calculate as shown in Table 2 with an assumed chilling weight loss of 1%. Table 2. Average heat load in a typical beef carcass chiller Sensible Heat Latent Heat [kw]/[btu/h] [kw]/[btu/h] Conduction 5.4 (18,442) 0 Infiltration 0 0 Product 35.6 (121,580) 9.3 (31,761) Fans 6.6 (22,540) 0 Total 47.6 (162,563) 9.3 (31,761) 14 IIAR 2009 Technical Paper #7

17 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective In Table 3 is shown a comparison of estimated chilling weight loss between two air cooler types which have identical overall surface areas. Both coolers are shown in end view (Figure 19) and both service the carcass chiller which is the subject of the load calculation (Table 2). One cooler is squat; the other has a relatively large face area and is shallow in the direction of air flow. Table 3. Estimated theoretical chilling weight loss for two different air cooler geometries and identical total heat exchanger surface areas. Number of coils per chiller 2 3 Estimated chilling weight loss, kg (lbs) 480 (1058) 270 (595) Chilling weight loss, % Value of loss, $/p.a. (1999 values) 240, ,000 Approximate energy costs, $/p.a. (1999 values) 10,300 12,200 The higher energy cost for the low weight loss design is due to the greater circulated air quantity. The difference in equipment costs is around $30,000 in 1999 values with the low weight loss solution being the most expensive. Provided the operator gets paid for the achieved weight gain, this differential investment may be returned in a few months. Discussion Correct technical selection of an extended surface air cooler for an industrial refrigeration application is a complex matter for the refrigeration contractor. It requires careful evaluation of a large number of issues ranging from cooling capacity, total/sensible heat ratios, relative humidity, water droplet carry-over, defrost frequency, sound levels and air throw to oil drainage, defrost methods, operating weight, material suitability, hygiene, corrosion resistance, delivery time and capital costs. If the technical selection process being undertaken by the refrigeration Technical Paper #7 IIAR

18 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas contractor is being complicated further by equipment manufacturers attempting to influence purchasing decisions by way of technical marketing material displaying varying degrees of relevance and indeed integrity, then it is not a surprise that the whole process can become confusing, at times overwhelming and prone to an easy way out approach. This paper has shown variations in heat transmission coefficients of around a factor of two as a direct result of changes to coil geometry only (i.e., variations in tube diameters, tube centers and tube patterns). In addition, a variation in heat transmission coefficients by a factor of almost three was shown as a direct result of changes in materials of construction and fin thickness for constant coil geometry. These two statements highlight the inherent danger of rules of thumb selecting air coolers on the basis of simple surface area allocation. An area of great significance in practice is refrigerant circuiting. The refrigerant circuiting of an air cooler is dictated by heat transfer fundamentals. Yet there are thousands of examples of air cooling coils performing poorly due to incorrect circuiting. This is particularly the case in dry expansion applications where optimal circuiting is only possible inside a relatively limited envelope of operating conditions. The key question for the refrigeration contractor is how is it possible to evaluate whether or not the potential supplier of a particular cooling coil is proposing the correct circuiting? The first step is to request information. There are many examples of the circuiting details not being disclosed by manufacturers at design stage. Once the equipment is on site or installed it is usually too late. Although the issue of fouling has been discussed, the issue of frost accumulation has not. Frost accumulation on the external surfaces of the cooling coil will inhibit air flow. A reduction in air flow will reduce cooling capacity and this has not been accounted for in Figure 12, which focuses on the thermal resistance only. This raises an important issue to consider for the refrigeration contractor. The issue is the 16 IIAR 2009 Technical Paper #7

19 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective allowance of sufficient frost volume on the external surfaces of the cooling coil particularly at coil inlet to enable continued operation without inhibiting air flow. It is commonly known that increasing the face velocity through a cooling coil improves heat transfer and coil capacity. What may not be readily identifiable is that a reduction in coil face velocity from 3.5 m/s (689 fpm) to 2.5 m/s (492 fpm) may only reduce gross coil capacity by 18% yet reduce fan power by 60% and in many cases leave net coil capacity (gross capacity minus fan power) almost unchanged. This represents two potential traps for the refrigeration contractor. Firstly, there is the likelihood of making a misdirected purchasing decision on the basis of gross coil capacity. Secondly, excess fan heat may be detrimental to some perishable products because the consequence will be a reduction in relative room humidity. The message for refrigeration contractors generally and particularly with respect to the selection of extended surface air coolers is to devote the topic the engineering attention it deserves. A well known slogan from the oil industry is oil ain t oils. This slogan is equally applicable to the refrigeration industry in the slightly modified form coils ain t coils. Technical Paper #7 IIAR

20 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas Nomenclature LMTD [K] Logarithmic Mean Temperature Difference ETD [K] Entering temperature difference m [kg/s] mass flow α [-] Orifice contraction coefficient A [m 2 ] Orifice area p [Pa] Pressure drop [kg/m 3 ] Density p [-] Number of tube passes t E [ C] Evaporating temperature t S [ C] Saturated suction temperature p R [K] Refrigerant pressure (temperature) drop d [m] Diameter H [m] Height FeZn [-] Mild steel, hot dipped galvanized SS [-] Stainless steel Al [-] Aluminum FIN [m] Fin thickness k O [W/m 2 K] Heat transmission coefficient u [Btu/h ft 2 F] Heat transmission coefficient Q E [kw] Total evaporator capacity 18 IIAR 2009 Technical Paper #7

21 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Q E,SENS [kw] Gross sensible evaporator capacity Q 0 [kw] Net sensible evaporator capacity Q FAN [kw] Fan heat rejection i [W/m 2 K] Inside film coefficient W FACE [m/s] Face velocity n FAN [-] Rotational fan speed Technical Paper #7 IIAR

22 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas References [1] Jensen, S.S Extended Surface Steel Air Coolers for Industrial Refrigeration, AIRAH Journal August 1989: [2] Jensen S.S Dry Expansion Feed in Dual Stage Ammonia Plants: Operating Experiences in a Large Refrigerated Distribution Centre. Proc IIAR Ammonia Refrigeration Conference and Exhibition, Reno, Nevada. [3] Koster, G.J Grenco, s Hertogenbosch, Netherlands. Energy Savings in Ammonia Refrigeration Plant by Using Oil Scrubbers. Proc. Institute of Refrigeration [4] Mälhammar, Å Frost Growth in Evaporators. Scandinavian Refrigeration 6/86: [5] Jensen, S.S Design and Selection of Industrial Finned Air Coolers for Natural Refrigerants A Comparison between NH 3 and CO 2, Proc. The Natural Refrigerants Transition Board Design Seminar, Sydney, Australia. [6] Jensen S.S Carcass Chilling After Slaughter, Innovative Air Cooling and Air Distribution Techniques. Proc. Australian Meat Council CEO Conference. 20 IIAR 2009 Technical Paper #7

23 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Figure 1. Typical extended surface stainless steel/aluminum air cooler Technical Paper #7 IIAR

24 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas Figure 2. Impact of cooler geometry on heat transmission coefficient (u-value) Figure 3. Impact of materials of construction and fin thickness on heat transmission coefficient 22 IIAR 2009 Technical Paper #7

25 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Figure 4. Common circuiting method with horizontal headers and bottom refrigerant inlet for industrial liquid overfeed applications Figure 5. Multiple header arrangement Technical Paper #7 IIAR

26 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas Figure 6. Circuit length reduction by employing vertical headers, coil shown in plan view Figure 7. Equilibrium refrigerant flows through the circuits of a cross flow finned air cooler, high heat flux, ammonia refrigerant, liquid overfeed 24 IIAR 2009 Technical Paper #7

27 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Figure 8. Impact of refrigerant circuiting on coil performance Figure 9. Physical parallel flow circuiting for thermodynamic counter flow Technical Paper #7 IIAR

28 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas Figure 10. Coil circuiting with parallel flow and reversed suction return Figure 11. Performance comparison for large low temperature finned air cooler with physical parallel flow and physical counter flow air/refrigerant flow patterns 26 IIAR 2009 Technical Paper #7

29 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Figure 12. Relative impact of fouling on coil performance; f i represents inside fouling, f o outside fouling Technical Paper #7 IIAR

30 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas Figure 13. Coil performance as a function of liquid overfeed rate 28 IIAR 2009 Technical Paper #7

31 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Figure 14. Evaporator with gravity flooded refrigerant feed Figure 15. Capacity of an NH 3 evaporator as a function of entering temperature difference and face velocity. Dry expansion (DX) and liquid overfeed (LR), air on constant at 35 C ( 31F) Technical Paper #7 IIAR

32 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas Figure 16. Air pressure drops for different coil geometries; coil layouts and operating conditions as per appendix 1 Figure 17. Energy optimization at varying face velocities and entering temperature differences 30 IIAR 2009 Technical Paper #7

33 Extended Surface Air Coolers for Industrial Plants the Contractors Perspective Figure 18. Commonly used batch carcass chiller layout in side elevation. Figure 19. End view of the two different evaporator geometries and layouts used for the weight loss estimate provided in Table 3 Technical Paper #7 IIAR

34 2009 IIAR Industrial Refrigeration Conference & Exhibition, Dallas, Texas Notes: 32 IIAR 2009 Technical Paper #7